During rotation of an axial flow rotary valve, openings in the periphery of the valve are arranged to periodically communicate with a similar window in the bore of the cylinder head that opens directly into the combustion chamber. Alignment between the openings and the window allows the passage of gas from the valve to the combustion chamber or vice versa. During a portion of the cycle when compression and combustion of gases takes place, the periphery of the valve blocks the window. The gas sealing system prevents the escape of high pressure gas during this portion of the cycle.
The valve is typically supported by bearings located either side of a centre portion of the valve in which the openings in the valve's periphery are located. The valve and its bearings are housed in a bore in the cylinder head in such a fashion as to ensure the centre portion can rotate whilst always maintaining a small radial clearance to the bore.
The bearings are lubricated with oil, and in some instances the valve is also cooled with oil that is pumped through the valve. An oil sealing system prevents the migration of oil into the area between the periphery of the centre portion of the valve and the bore in which the valve is housed.
Large numbers of rotary valve arrangements have been proposed but none have achieved commercial success. One of the major contributing factors to this lack of success is the failure to develop a satisfactory gas sealing arrangement.
Rotary valves that run with a small predetermined clearance to the cylinder head bore were developed to address the problem of thermal and mechanical deflection of the valve that had plagued rotary valve development for the most part of the 20th century. One of the first of these solutions is found in U.S. Pat. No. 4,852,532 (Bishop), in which the valve is designed to run with a small clearance to cylinder head bore and the combustion chamber is sealed by means of a floating array of seals arranged around the window in the cylinder head.
The existence of a small clearance between the periphery of the valve and the cylinder head bore enabled a small amount of thermal and mechanical distortion to occur without the periphery of the valve making contact with the bore. However, in valves with both inlet and exhaust ports in the same valve this means that these ports were not physically sealed from one another.
In these arrangements it is necessary to control the flow of gas between the ports—either the flow of exhaust gas from the exhaust port to the inlet port or the flow of inlet gas from the inlet port to the exhaust port. Flow of gas between the inlet and exhaust port was controlled by making the radial clearance between the valve and the cylinder head bore small. The combination of small clearances and small pressure differences between the inlet and exhaust ports limits the flow of gases between the ports. In most cases the pressure in the exhaust port is larger than the pressure in the inlet port, and consequently leakage is predominantly from the exhaust port into the inlet port. This sealing arrangement requires the use of valves that have minimal thermal and physical distortion. The smaller the distortion, the smaller the required radial clearance and the better the sealing between the inlet and exhaust ports.
In some instances the flow of a limited amount of gas between the ports is acceptable. In most engines a small amount of flow from the exhaust port into the inlet port is not a great concern as it merely contributes to a process that is often used to regulate emissions—ie internal egr (exhaust gas recirculation). However, in some instances such as high performance engines even a small amount of leakage between the exhaust port and the inlet port is a problem as this leakage reduces engine power.
Pressures in the exhaust and inlet ports are not constant over the duration of an engine cycle. The pressures in these ports undergo large cyclic variations during the engine cycle. For example in some engines, shortly after the exhaust opens, the exhaust port pressure rises rapidly above atmospheric pressure to over 1 bar. Later in the cycle the pressure may fall below atmospheric pressure. In high performance engines at full throttle and when the exhaust valve is closed, pressures in the exhaust port will vary sinusoidally as pressure waves traverse between the exhaust opening and the end of the exhaust pipe.
In the inlet port, the pressure will be pulled low early in the induction process and will rise above atmospheric pressure later in the induction process close to the point where the valve closes. When the inlet opening is closed pressure in the inlet port will vary sinusoidally as pressure waves traverse between the inlet opening and the end of the inlet tract. As a consequence, the pressure difference between the inlet and exhaust port is subject to rapid cyclic fluctuation through out the engine cycle.
In other instances, such as a production engine operating at idle, the amount of leakage between the exhaust port and the inlet port may be too great and adversely affect the smooth idle of the engine. In production engines having a throttle operating at idle, the pressure in the inlet port is maintained at a very low pressure, typically lower than 0.5 bar below atmospheric pressure (−0.5 bar gauge), over most of the engine cycle. The low engine speed at idle means there is a relatively long period of time for exhaust to leak across the bridge into the inlet port. The combination of high pressure difference and long time available for leakage means that in some instances excessive amounts of exhaust gas will leak into the inlet system.
The purpose of the present invention is to provide a mechanism that minimises the flow of gas from one port to another, and in particular during those portions of the cycle where there is a large pressure difference between the ports and where the resulting flows will create problems with the operation of the engine.
A typical example of where this is issue occurs is in high performance engines under full throttle. Because of the nature of these engines, they require larger radial clearances and are therefore more susceptible to adverse pressure drops between the ports. Under full throttle operation shortly after the exhaust port first opens the pressure in the exhaust port rises rapidly to well over 1 bar gauge. In arrangements like those in U.S. Pat. No. 4,852,532 (Bishop) this will result in the exhaust being pushed into the inlet port and a consequent loss of performance. Later in the cycle when the exhaust pressure drops below atmospheric, the transfer of exhaust gas to the inlet port is no longer a problem.
Another example occurs in engines operating under closed throttle conditions. During the induction stroke the cylinder and the inlet port are both pulled down to a very low pressure. If this occurs after the bridge between the inlet and exhaust port has passed over the trailing axial seal there is direct communication between the cylinder, the inlet port and the exhaust port through the gap that exists between the bridge and the bore in which the valve is housed. Exhaust will flow across the bridge into the inlet port driven by this high pressure drop between the ports.
The present invention is designed to work with a gas sealing system that consists of an array of floating seals positioned around the window and an oil sealing arrangement, both of which are disclosed in International Application PCT/AU2005/001306 published as WO 2006/024081 A1 (Bishop Innovation Limited) having the same priority date and applicant as the present invention. The present invention can however be adapted to work with other gas sealing arrangements having an array of floating seals such as the seal array disclosed in U.S. Pat. No. 5,509,386 (Wallis et al).
The closest known prior art is U.S. Pat. No. 5,941,206 (Smith et al). This patent applies to a different class of rotary valve than the present invention. However, the patent purports to solve the same problem that the present invention address, ie the problem of leakage between the ports in a rotary valve that maintains a clearance to the cylinder bore in which the valve is housed. It purports to solve this problem by fitting axial extending seals into the rotating valve and loading them against the cylinder head bore in which the valve is housed. In this arrangement axial extending seals are positioned between the ports thereby preventing any significant transfer of gas circumferentially directly between the ports irrespective of the pressure difference between the ports. However for reasons explained below it fails to prevent the transfer of gas between the ports and worse still it allows the combustion gases from the cylinder to leak into the ports and in the process fails to adequately seal the combustion chamber. Furthermore the type of arrangement disclosed has several other problems as follows.
Firstly, consider the sealing system. During combustion the high pressure trapped between an adjacent pair of axial extending seals and the circumferential seals will push the circumferential seals axially outward against the axially outer faces of the grooves in which the circumferential seals are housed. High pressure gas will be forced into the gap between the axially inner face of groove and the circumferential seal and from there under the circumferential seal. This high pressure gas sitting in a L shaped clearance volume formed between the circumferential groove and the circumferential seal is free to fill the entire circumference of this L shaped clearance volume. The high pressure gas in this L shaped clearance volume is free to exit axially inward at all positions around the circumference circumferentially outboard of the two axial extending seals that straddle the window. No matter how much care is made to make the clearance between the circumferential seal and its groove small the leakage areas are huge due to the long circumferential periphery outside the axial extending seals. Worse still the high pressure gas forced under the axial extending seals is able to discharge directly into the L shaped clearance volume. The high pressure combustion gases in this L shaped clearance volume will discharge into the ports. A similar problem was encountered on earlier axial flow rotary valve arrangements and this was addressed in U.S. Pat. No. 5,526,780 (Wallis). Furthermore, gas in the ports will be able to leak between the ports using the same mechanism.
Secondly, housing the axial extending seals in the rotary valve unnecessarily increases the diameter of the valve. The valve diameter is a critical issue on rotary valve engines as it is a dominant determinant as to where the spark plug can be positioned. The larger the valve diameter the further the plug must be placed from the theoretical optimum position in the centre of the cylinder. The engine of U.S. Pat. No. 5,941,206 (Smith et al) clearly demonstrates this problem. It has a valve with a diameter similar to that of the cylinder bore and the spark plug is located under the valve and well away from the centre of the cylinder. This arrangement is acceptable for a multicylinder engine, as the valve is parallel to the crankshaft and the spark plugs can be positioned along the outside of the engine rather than between the cylinders. However, it would have no application on an axial flow rotary valve of the present invention where there is a single valve per cylinder and the axis of the valve is generally perpendicular to the crankshaft axis. In these arrangements the spark plug has to be located between the cylinders.
Thirdly, the axial seals rotate with the valve and relative to a stationary window. Consequently the distance of the seals from the sides of the window varies as the valve rotates. Whenever the seals are not immediately adjacent the window there will be excessive crevice volume that will adversely affect performance and fuel efficiency. The crevice volume issues are addressed in U.S. Pat. No. 5,526,780 (Wallis). Furthermore some of the axial extending seals must pass through the combustion chamber during the combustion process exposing the seal to unnecessarily high temperatures and heat flows.
Finally, as the axial seals rotate with the valve they seal and rub against the cylinder head bore. Consequently the cylinder head bore must be hard to prevent wear and scuffing. This clearly limits the types of material that the head can be made from. Cylinder heads are most commonly made from aluminium. If this material was used it would have to be hard coated adding to expense and complication. Alternatively as disclosed in U.S. Pat. No. 5,941,206 (Smith et al) an additional sleeve must be fitted to the head increasing the complexity and cost.
The present invention seeks to ameliorate one or more of the disadvantages associated with the abovementioned prior art.